Hydraulic control system of an automatic power transmission

ABSTRACT

An improved hydraulic control system is provided for use in an automatic power transmission for an automotive vehicle, comprising shift valves at least one of which has hysteresis characteristics which are reduced during kick-down condition so that the shift valve is moved from the upshift to downshift position at a vehicle speed which is higher than a level at which the shift valve of a prior art control system is moved from the upshift to downshift position.

United States Patent Mitamura et al.

HYDRAULIC CONTROL SYSTEM OF AN AUTOMATIC POWER TRANSMISSION Inventors: Kenichi Mitamura; Kunio Ohtsuka.

both of Yokohama; Toshiyuki Miyauchi, Kamakura, all of Japan Assignee: Nissan Motor Company Limited,

Yokohama, Japan Filed: Aug. 30, 1974 Appl. No.: 502.299

Foreign Application Priority Data Oct. I8, 1973 Japan 48-116297 u.s. Cl. 74/869; 74/864; 74/868; 74/752 C 1m. cit 860K 41/06; B60K 41/18; F16I-1 3/74 Field of Search 74/863, 864, 867, 868, 74/869, 752 c References Cited UNlTED STATES PATENTS 6/1972 lrie 74/752 C FRONT CLUTCH REAR CLUTCH LOW- REV.

BRAKE 1 1 Sept. 30, 1975 Nagamatsu 74/869 3.670.599 6/1972 3.685.372 8/1972 Miyazaki 74/869 X 3.859.873 1/1975 Miyauchi et al. 74/867 FOREIGN PATENTS OR APPLICATIONS 1,923,577 11/1969 Germany 74/863 Primary E.\'zuniner-Samuel Scott Assistant Examiner-Lance W. Chandler [57] ABSTRACT An improved hydraulic control system is provided for use in an automatic power transmission for an automotive vehicle, comprising shift valves at least one of which has hysteresis characteristics which are reduced during kick-down-condition so that the shift valve is moved from the u'pshift to downshift position at a vehicle speed which is higher than a level at which the shift valve of a prior art control system is moved from the upshitt to downshift position.

5 Claims, 7 Drawing Figures 2ND- SPEED SPEED DRIVING BRAKE US. Patent Se t. 30,1975 ghee t 1 0 3,908,486

Fig-

US. Patent Sept. 30,1975 Sheet 2 of6 3 9%,486

Fig. 2a

BAND BRAKE APPLY RELEASE SIDE SIDE US. Patent Sept. 30,1975 Sheet 3 of6 E 3,908,486

Fig. 2b

2ND- FRONT REAR LOW- REV. SPEED SPEED CLUTCH CLUTCH BRAKE COASTING DRI US. Patent Sept. 30,1975 Sheet4 0'56 $998,486

Fig. 2C

US. Patent Sept. 30,1975 Sheet 5 of6 3,908,486

Fig. 20'

US. Patent Sept. 30,1975 Sheet 6 of6 3,99,486

MANIFOLD VACUUM 550mm Hg mmiwwmml m tbm I.

oz ZmEO mHEIOKIP VEHICLE SPEED HYDRAULIC CONTROL SYSTEM OF AN AUTOMATIC POWER TRANSMISSION This present invention relates to automatic power transmissions of automotive vehicles and, more particularly, to a hydraulic control system for use in the automatic power transmissions.

The hydraulic control system of the automatic power transmission has incorporated therein a shift valve to effect downshift or upshift between gear ratios during automatic drive condition of the transmission. The shift valve is acted upon by a line pressure and a throttle pressure which is produced from the line pressure in relation to variation in engine load. The shift valve is usually provided with differential pressure-acting areas by means of which a greater force urges the shift valve into the downshift position when the shift valve is in the downshift position than a force which urges the shift valve toward the downshift position when the shift valve is in the upshift position. The shift valve is therefore moved from the upshift position to the downshift position at a vehicle speed which is lower than the vehicle speed at which the shift valve is moved from the downshift position to'the upshift position. Such a differential action of the shift valve is herein called the hysteresis characteristics of the shift valve. The line pressure and the throttle pressure increase as the throttle valve of the carburetor is moved toward the full-throttle position as is well known in the art and, for this reason, the hysteresis characteristics of the shift valve tend to be amplified as the carburetor throttle valve approaches the full-throttle position. The hysteresis characteristics of the shift valve thus peak up when the accelerator pedal is depressed all the way down to produce a kickdown condition in the engine. If, thus, the hysteresis characteristics of the shift valve are so determined as to provide appropriate downshift and upshift patterns during usual operational conditions, viz., nonkickdown conditions of the engine and to effect the upshift at a preferred vehicle speed during kick-down condition, then the shift valve will be moved from the upshift to downshift position at a vehicle speed which is lower than a desirable level.

It is, therefore, an object of the present invention to provide an improved hydraulic control system of an automatic power transmission featuring a shift valve which is bestowed with such hysteresis characteristics as are not only appropriate during the usual or nonkickdown conditions but are reduced during kick-down condition.

In accordance with the present invention, such an object will be accomplished in a hydraulic control system which comprises a source of line pressure, a throttle valve responsive to load on an engine for producing from the line pressure a throttle pressure which is related to the engine load, a governor valve responsive to vehicle speed for producing a governor pressure which is related to the vehicle speed, a hysteresis valve for producing from the line pressure a substantially constant hysteresis pressure which is higher than the governor pressure, a kick-down valve communicating with the throttle valve and operative to pass therethrough the hysteresis pressure when the throttle pressure is increased to a level substantially representative of kickdown condition of the engine, and a plurality of shift valves each having a downshift position and an upshift position, at least one of the shift valves having first differential pressure-acting areas and second pressure acting areas and formed with a first fluid port to communicate with the source of the line pressure during automatic drive range condition of the transmission, the line pressure in the first fluid port acting upon the first differential pressure-acting areas for urging the shift valve into the downshift position when the shift valve is in the downshift position, the line pressure in the first fluid port being isolated from the first differential pressure-acting areas when the shift valve is in the upshift position, a second fluid port which is in constant communication with the throttle valve, the throttle pressure in the second fluid port urging the shift valve toward the downshift position, a third fluid port communicating with the governor valve, the governor pressure in the third fluid port urging the shift valve toward the upshift position, a fourth fluid port in constant communication with the hysteresis valve and located to be closed when the shift valve is in the upshift position, the hysteresis pressure in the fourth fluid port acting upon the second differential pressure-acting areas for urging the shift valve into the downshift position when the shift valve is in the downshift position, and a fifth fluid port communicating with the kick-down valve and located to be closed when the shift valve is in the downshift position, the hysteresis pressure passed through the kick-down valve to the fifth fluid port acting upon the second differential pressuare-acting areas for urging the shift valve toward the downshift position when the shift valve is in the upshift position and when the kickdown valve is open.

The hydraulic control system may further comprise a control valve which is responsive to kick-down condition of the engine and connected between the hysteresis valve and the kick down valve for being open to pass therethrough the hysteresis pressure from the hysteresis valve to the kick-down valve in response to the kickdown condition of the engine. The control valve may be modified so as to be responsive not only to the kickdown condition but an idling or part-throttle condition of the engine. In this instance, the hydraulic control system may still further comprise an idle valve which is connected between the throttle valve and the hysteresis valve for being open to pass therethrough the hysteresis pressure of a level which is substantially representative of idling condition of the engine, and a two-position valve which has a first inlet port communicating with the governor valve, a second inlet port communicating with the idle valve and an outlet port communicating with the third fluid port of the shift valve, the twoposition valve being operative to have the first inlet port open to pass the governor pressure from the governor valve to the third fluid port of the shift valve or to have the second inlet port open to pass the hysteresis pressure from the idle valve to the third fluid port of the shift valve when the control valve and the idle valve are concurrently open.

In this instance, the hydraulic control system may further comprise a downshift valve which is connected between the governor and the first inlet port of the twoposition valve and which is in communication with the kick-down valve, wherein the downshift valve is operative to boost the governor pressure from the governor valve when the kick-down valve is open so that the hysteresis pressure is directed to the downshift valve through the control valve and the kick-down valve.

With the hydraulic control system thus constructed and arranged, the shift valve is urged into the downshift position by the line pressure acting on the differential pressure-acting areas, the hysteresis pressure acting upon the second differential pressure-acting areas and the throttle pressuare when the shift valve is in the downshift position irrespective of the operational conditions of the engine. When, however, the shift valve is in the upshift position, the shift valve is urged toward the downshift position only by the throttle pressure during usual or non-kickdown condition and by the throttle pressure and the hysteresis pressure acting on the second differential pressure-acting areas. The hysteresis characteristics of the shift valve are thus provided by the line pressure acting on the first differential pressure-acting areas and the hysteresis pressure acting on the second differential pressure-acting areas when the shift valve is in the downshift position during usual or non-kickdown condition of the engine. During the kick-down condition, however, the hysteresis pressure of the shift valve result only from the line pressure acting on the first differential pressure-acting areas of the shift valve in the downshift position. The shift valve is, in this manner, moved from the upshift position to the downshift position at a vehicle speed which is higher than the levels at which the shift valves in the prior art hydraulic control system of the automatic power transmissions are moved from the upshift to the downshift positions during the kickdown conditions in the automatic drive range of the transmissions. Where the previously mentioned downshift valve is incorporated into the hydraulic control system according to the present invention, the governor pressure to be directed to the shift valve is boosted when the kick-down valve is closed responsive to the usual or non-kickdown condition of the engine or, in other words, the governor pressure is passed without being boosted by the downshift valve to the shift valve when the kickdown valve is open to pass therethrough the hysteresis pressure. During the kick-down condition of the engine, therefore, the shift valve will be moved form the upshift position to the downshift position or vice versa at a higher vehicle speed than during the usual operational condition of the engine.

Other features of the hydraulic control system according to the present invention will become more apparent from the following description taken in conjunction with the accompanying drawings in which:

FIG. 1 is a schematic view which shows an example of the automatic power transmission mechanism with which the hydraulic control system according to the present invention may be combined;

FIGS. 2a-2d is a schematic view which shows a pre ferred embodiment of the hydrualic control system according to the present invention;

FIG. 3 is a graph which is representative of a relation between the vacuum in the intake manifold of an engine and the throttle pressure produced in the hydraulic control system illustrated in FIG. 2; and

FIG. 4 is a graph which indicates shift patterns achieved in the prior art hydraulic control system and in the control system iullustrated in FIG. 2.

Reference will now be made to the drawings, first to FIG. 1 which shows an example of the automatic power transmission mechanism to which the hydraulic control system embodying the present invention is applicable. The automatic power transmission mechanism is herein shown as being of the four-forward-speed and onereverse-speed type. This is, however, merely by way of example and, as such, the hydraulic control system according to the present invention may be incorporated into any of other types of automatic power transmission such as for example a three-forward-speed and onereversespeed transmission or a five-forward-speed and one-reverse-speed transmission.

The automatic power transmission mechanism is shown to largely consist of a transmission case which is generally designated by reference numeral 10, a torque converter 12, an oil pump 14, a transmission input shaft 16, a first or front clutch 18, a second or rear clutch 20, first, second and third planetary gear sets 22, 24 and 26, respectively, a low-and-reverse brake 28, a secondspeed-coasting brake 30, a second-speed-driving brake 32, a band brake 34, a low-one-way clutch 36, a second-speed one-way clutch 38 and a transmission output shaft 40. The torque converter 12 is assumed to be of the three-member type and thus consists of a driving member or impeller 42, a driven member or turbine 44, and a reaction member or stator 46 as is customary in the art. The impeller 42 is connected through a torque converter drive plate 48 to the crankshaft 50 of a power plant such as an internal combustion engine 52 and is thus driven to roatate at engine speed when the engine is in operation. The turbine 44 is mounted on, or usually splined to, the transmission input shaft 16 and drives the transmission input shaft 16 through speed and torque ratio changes when the impeller 42 is driven by the engine 52. The stator 46 is connected to the transmission case 10 through a one-way clutch 54. The stator 46 is thus allowed to rotate in the same direction as the impeller 42 and is prevented from being rotated in the opposite direction. The impeller 42 is further connected through an impeller shaft 56 to the oil pump 14. The oil pump 14 is thus operative to deliver a working fluid under pressure when the impeller 42 of the torque converter 12 is driven by the engine 48. The transmission input shaft 16 is connected to clutch discs 18a of the front clutch l8 and to clutch plates 20a of the rear clutch 20 through a common connecting member 58.

To first, second and third planetary gear sets 22, 24 and 26 consist respectively of externally toothed sun gears 22a, 24a and 26a rotatable about their respective axes which are in line with the transmission output shaft 40, planet pinions 22b, 24b and 26b which are respectively in mesh with the sun gears 22a, 24a and 26a and which are rotatable about their respective axes around the mating sun gears, and internally toothed ring gears 22c, 24c and 260 which are respectively in mesh with the planet pinions 22b, 24b and 26b about the axes of the associated sun gears. The sun gears 22a and 24a of the first and second planetary gear sets 22 and 24, respectively, are connected to clutch plates 18b of the front clutch 18 through a connecting drum 60. When, thus, the clutch discs 18a and the clutch plates 18b are engaged by each other, the front clutch 18 is coupled so as to transmit the driving power from the transmission input shaft 16 to the connecting drum so that the sun gears 22a and 24a of the first and second planetary gear sets 22 and 24, respectively, are driven from the engine 52. The connecting drum 60 is wrapped with the band brake 34. When, thus, the brake band 34 is actuated to be tightened upon the drum 60, the drum 60 is locked and accordingly the sun gear 22a of the first planetary gear set 22 and the sun gear 24a of the second planetary gear set 24 are held stationary. The planet pinions 22b of the first planetary gear set 22 are carried by a pinion carrier 62 and are connected through the pinion carrier 62 to the transmission output shaft 40. The ring gear 220 of the first planetary gear set 22 is connected to clutch discs 20b of the rear clutch 20 through an intermediate shaft 64. which is in line with the transmission input and output shafts l6 and 40. When the clutch plates 20a and the clutch discs 20b are engaged by each other, the rear clutch 20 transmits the driving power from the transmission input shaft 16 to the intermediate shaft 64 so that the ring gear 220 of the first planetary gear set 22 is driven from the engine 52 through the torque converter 12. The planet pinions 24b of the second planetary gear set 24 are carried by a pinion carrier 66 and likewise the planet pinions 26b of the third planetary gear set 24 are carried by a pinion carrier 68. The pinion carriers 66 and 68 are connected together by a connecting drum 70 which is positioned is surrounding relation to the transmission output shaft 40. The ring gear 24c of the second planetary gear set 24 and the sun gear 26a of the third planetary gear set 26 are connected to the transmission output shaft 40 through connecting members 72 and 74, respectively.

The low-and-reverse brake 28 has stationary brake plates 28a fixed to the transmission case and movable brake discs 28b connected to the connecting drum 70 which interconnects the pinion carriers 66 and 68 of the planet pinions 24b and 26b of the second and third planetary gear sets 24 and 26, respectively. When, thus, the movable brake discs 28b are actuated into engagement with the stationary brake plates 28a, the lowand-reverse brake 28 becomes operative to brake the connecting drum 70 so that the planet pinions 24b of the second planetary gear set 24 and the planet pinions 26b of the third planetary gear set 26 are held stationary. The low-and-reverse brake 28 is paralleled by the low-one-way clutch 36 which has an outer race member 36a fixed to the transmission case 10 and an inner race member 36b connected to the connecting drum 70. The low-one-way clutch 36 is coupled to transmit a driving torque from the transmission input shaft 16 to the transmission output shaft 40 and is uncoupled to idle away (with the inner race member rotating freely in the outer race member) when a driving torque is imparted to the one-way clutch 36 from the transmission output shaft 40.

The second-speed-coasting brake 30 has stationary brake plates 30a fixed to the transmission case 10 and movable brake discs 30b connected to the ring gear 260 of the third planetary gear set 26 through a connecting drum 76. When, thus, the movable brake discs 30b are actuated to be engaged by the stationary brake plates 300, the second-speed-coasting brake 30 becomes operative to brake the connecting drum 76 so that the ring gear 26c of the third planetary gear set 26 is held stationary, The connecting drum 76 is further connected to the sccondspeed-driving brake 32 through the second-speed one-way clutch 38. The secondspeed-driving brake 32 has stationary brake plates 32a which are fixed to the transmission case 10 and movable brake discs 32b which are connected to an outer race member 38a of the second-speed one-way clutch 38 which has an inner race member 38b connected to the connecting drum 76. The second-speed one-way clutch 38 is adapted to be coupled when driven by a driving torque transmitted thereto from the transmission input shaft 16 and is uncoupled when a driving torque is imparted thereto from the transmission output shaft 40. When, thus, a driving torque is being transmitted from the transmission input shaft 16 to the second-speed one-way clutch 38 and at the same time the movable brake discs 32b of the second-speeddriving brake 32 are engaged by the stationary brake plates 32a, both of the second-speed-driving brake 32 and the second-speed one-way clutch 38 are coupled so that the connecting drum 76 is braked upon by the second-speed-driving brake 32 and accordingly the ring gear 26 c of the third planetary gear set 28 is held stationary. When a driving torque is transmitted to the second-speed one-way clutch 38 from the transmission output shaft 40, then the inner race member 38b of the one-way clutch 38 freely rotates in the outer race member 38a so that the driving torque from the output shaft 40 is not transmitted to the ring gear 260 of the third planetary gear set 26 with the second-speed one-way clutch 38 idling away. Designated by reference numeral 78 is a governor unit which is mounted on the transmission output shaft 40 for producing a fluid pressure which varies with the revolution speed of the transmission output shaft 40 as will be in more detail discussed later. The low-one-way clutch 36 and the second-speed one-way clutch 38 will have been described as being of the Sprag type. This is, however, merely by way of example and as such the clutches 36 and 38 may be of the known cam and roller type where desired.

The front clutch 18, the rear clutch 20, the low-andreverse brake 28, the second-speed-coasting brake 30, the secondspeed-driving brake 32 and the band brake 34 are actuated and the low-one-way clutch 36 and the second-speed one-way clutch 38 are coupled selectively in accordance with schedules which are indicated in Table 1, wherein a symbol 0 is indicative of the coupled condition of each of the clutches 18 and 20 and the brakes 28, 30, 32 and 34 and a symbol x is indicative of the coupled condition of each of the oneway clutches 36 and 38. The gear ratios indicated Table 1 Selected Front Rear Low-rev. 2nd-speed- 2nd speed- Band Low 2nd-speed- Gear speeds clutch clutch brake coasting driving brake oneway oneway ratios brake brake brake brake 18 20 28 30 32 34 36 38 Automatic 1st 0 x 2.4611

Drive 2nd 0 o x 2.00:1

Range 3rd 0 o 1.46:1

4th 0 o 1.00:1

Manual 1 st 0 o 246:]

Forward 2nd 0 o 2.00:1

Speed 3rd 0 0 1,4621 Range Reverse o o 2. 16:1

in the rightmost column of Table l have been calculated on the assumption that all the planetary gear sets 22, 24 and 26 have identical configurations and that the sun gear and ring gear of each of the planetary gear sets have 37 and 80 teeth, respectively. It is selfexplanetary that the gear ratios may be changed if the numbers of teeth of the sun and ring gears of the individual planetary gear sets are otherwise selected.

As will be seen from Table l, the conditions providing the first and second speeds in the automatic drive range differ from those providing the first and second speeds in the manual forward-speed range in that, while the driving torque is transmitted only in one direction from the input shaft 16 to the output shaft 40 during the conditions in which the first or second speed is established in the automatic drive range, the driving torque can be transmitted either from the input shaft 16 to the output shaft 40 or from the output shaft 40 to the input shaft 16 during the conditions in which the first or second speed is selected in the manual forwardspeed range. As previously mentioned, the low-oneway clutch 36 and the second-speed one-way clutch 38 are adapted to be coupled when given a driving torque from the transmission input shaft 16 and are uncoupled when subjected to a driving torque from the transmission output shaft 40. When, thus, the first or second speed is established in the automatic drive range with the low-one-way clutch 36 or the second-speed oneway clutch 38 held in a coupled condition, no driving torque can be transmitted from the transmission output shaft 40 to the transmission input shaft 16. This means that the engine 52 is unable to act as a brake when the vehicle is to be driven by an inertia of the vehicle if the first or second speed gear ratio is being attained during the automatic drive range. If it is desired to enable the engine to act as a brake as when, for example, the vehi cle is descending a hill, it is thus necessary to have the low-and-reverse brake 28 actuated to attain the firstspeed gear ratio and to have the second-speed-coasting brake 30 actuated to attain the second-speed gear ratio.

The operation of the automatic power transmission mechanism thus arranged will not be described.

To achieve the first-speed gear ratio in the manual forward-speed range or in the automatic drive range, the rear clutch 20 is coupled and simultaneously either of the low-and-reverse brake 28 and the low-one-way clutch 36 is actuated. In this instance, the motions taking place in the transmission mechanism will be easily understood if it is assumed that the driving torque originates in the transmission output shaft 40 and is transmitted from the transmission output shaft 40 to the transmission input shaft 16, the case being however actually to the contrary. When, thus, the transmission output shaft is rotated with the low-and-reverse brake 28 or the low-one-way clutch 36 coupled, the ring gear 240 of the second planetary gear set 24 is driven to rotate at the same speed and in the same direction as the transmission output shaft 40 through the connecting member 72 and concurrently the connecting drum 70 interconnecting the pinion carriers 66 and 68 of the second and third planetary gear sets 24 and 26, respectively, is braked by the low-and-reverse brake 28 or the low-one-way clutch 36. The sun gear 24a of the second planetary gear set 24 is consequently driven to rotate at a certain speed and in an opposite direction to the rotation of the ring gear 240 with the planet pinions 24b of the gear set 24 held stationary. The sun gear 22a of the first planetary gear set 22 is accordingly driven to rotate at the same speed and in the same direction as the sun gear 24a of the second planetary gear set 24. The pinion carrier 62 of the first planetary gear set 22 being connected to and rotating with the transmission output shaft 40, the ring gear 22c of the first planetary gear set 22 is driven to rotate in an opposite direction to the sun gear 22a (viz., in the same direction as the ring gear 24c of the second planetary gear set 24) and at a speed which is dictated by the revolution speeds of the sun gear 22a and the pinion carrier 62 of the planetary gear set 22. A driving torque is thus transmitted from the ring gear 220 of the first planetary gear set 22 to the transmission input shaft 16 through the intermediate shaft 64 and the rear clutch 20. Actually, the driving torque is transmitted in a reverse direction from the transmission input shaft 16 to the transmission output shaft 40, but the relative motions between the mating gears and the ratio between the revolution speeds of the transmission input and output shafts 16 and 40 are virtually similar to those above described. The first-speed gear ratio is in this manner determined by the numbers of teeth of the sun and ring gears of the first and second planetary gear sets 22 and 24.

To achieve the second-speed gear ratio, the rear clutch 20 is coupled and at the same time either the second-speed-coasting brake 30 is coupled for the manual forward-speed range or the second-speed-driving brake 32 is coupled for the automatic drive range. In this instance, it is also assumed for ease of understanding that the driving torque is transmitted from the transmission output shaft 40 to the transmission input shaft 16, contrary to the actual motions of the transmission mechanism. When, thus, the transmission output shaft 40 is rotated with the connecting drum 76 braked by either the second-speed-coasting brake 30 or the second-speed-driving brake 32, the sun gear 26a of the third planetary gear set 26 is driven at the same speed and in the same direction as the transmission output shaft 40 through the connecting member 74 and concurrently the ring gear 260 of the third planetary gear set 26 is held stationary. The pinion carrier 68 carrying the planet pinions 26b of the third planetary gear set 26 is consequently driven to rotate at a certain speed. Since the pinion carrier 68 of the third planetary gear set 26 is connected through the connecting drum 70 to the pinion carrier 66 carrying the planet pinions 24b of the second planetary gear set 24 and since the ring gear 240 of the second planetary gear set 24 is connected to the transmission output shaft 40 through the connecting member 72 and is thus driven to rotate at the same speed as the transmission output shaft 40, the sun gear 24a of the second planetary gear set 24 is driven to ro tate at a certain speed which is dictated by the revolution speeds of the pinion carrier 66 and the ring gear 240 of the second planetary gear set 24. Since, furthermore, the sun gear 24a of the second planetary gear set 24 is connected to the sun gear 22a of the first planetary gear set 22 and since the pinion carrier 62 carrying the planet pinions 22b of the first planetary gear set 22 is connected to the transmission output shaft 40 and is therefore driven to rotate at the same speed and in the same direction as the transmission output shaft 40, the ring gear 22c of the first planetary gear set 22 is driven to rotate in the same direction as the pinion carrier 62 and at a certain speed which is dictated by the revolution speeds of the sun gear 22a and the pinion carrier 62 of the first planetary gear set 22. The ring gear 220 of the first planetary gear set 22 being connected to the transmission input shaft 16 through the intermediate shaft 64 and the rear clutch 20, the transmission input shaft 16 is driven to rotate at the same speed and in the same direction as the ring gear 22c of the first planetary gear set 22. Although, in this instance, the actual direction of torque transmission between the transmission input shaft 16 and the transmission output shaft 40 is opposite to that above described, the relative motions between the mating gears and the ratio between the revolution speeds of the transmission input and output shafts l6 and 40 are virtually similar to those which have been described above.

For the purpose of achieving the third-speed gear ratio in the automatic drive range or in the manual for ward-speed range, the rear clutch is coupled and at the same time the band brake 34 is applied. The driving torque is consequently transmitted from the transmission input shaft 16 to the ring gear 22c of the first planetary gear set 22 through the rear clutch 20 and the intermediate shaft 64. Since, in this instance, the sun gear 22a of the first planetary gear set 22 is held stationary with the connecting drum 60 braked upon by the band brake 32, the pinion carrier 62 carrying the planet pinions 22b of the first planetary gear set 22 is driven to rotate at a certain speed and in the same direction as the ring gear 22c and drives the transmission output shaft 40 to rotate at the same speed as the pinion carrier 62.

The fourth-speed gear ratio of the automatic drive range can be produced when both of the front and rear clutches l8 and 20 are concurrently coupled. Under these conditions, the sun gear 22a and the ring gear 22c of the first planetary gear set 22 are driven to rotate at the same speeds as the revolution speed of the transmission input shaft 16 through the front clutch 18 and the rear clutch 20, respectively, and the first planetary gear set 22 is driven to rotate as a unit from the transmission input shaft 26 so that the transmission output shaft 40 is driven by the pinion carrier 62 of the first planetary gear set 22. Under the direct drive condition thus established, the transmission output shaft 40 is rotated at the same speed as the revolution speed of the transmission input shaft 16.

To achieve the reverse gear ratio, thefront clutch 18 is coupled and the low-and-reverse brake 28 applied. The driving torque is consequently transmitted from the transmission input shaft 16 to the sun gear 24a of the second planetary gear set 24 and at the same time the planet pinions 24b of the second planetary gear set 24 is held stationary because the connecting drum 70 interconnecting the pinion carriers 66 and 68 of the second and third planetary gear sets 24 and 26 is braked upon by means of the low-and-reverse brake 28. The ring gear 24c of the second planetary gear set 24 is consequently driven to rotate at a certain speed and in a direction opposite to the direction of rotation of the sun gear 24a of the planetary gear set 24. The driving torque from the transmission input shaft 16 is thus transmitted in a reverse direction to the transmission output shaft 40 through the ring gear 246 of the second planetary gear set 24.

The front clutch 18, the rear clutch 20, the lowandreverse brake 28, the second-speed-coasting brake 30, the second-speed-driving brake 32 and the band brake 34 are controlled in accordance with the previously described schedules by means of a hydraulic control system which is illustrated in FIG. 4. Referring to FIG. 4, the hydraulic control system comprises an oil pump 80, a manual selector valve 82, a vacuum-operated throttle valve 84, cooperating with a pressure modifier valve 86, a throttle back-up valve 88, a control pressure regu lator valve 90, a hysteresis valve 92, a solenoidoperated control valve 94, an idle valve 96, a kickdown valve 98, a governor valve unit which consists of a primary valve 102 and a secondary valve 104, a downshift valve 106, a first-second speed shift valve 108, a second-third speed shift valve 110, a third-fourth speed shift valve 112, and a second-speed pressure modulator valve 114, all of which valves are interconnected to one another by passageways cast in the trans mission case.

The pump 80 has a suction port connected through a fluid inlet passageway 116 and a strainer 1 18 to an oil reservoir 120. The pump 80, which is usually driven by the engine crankshaft, delivers a fluid under pressure from its delivery port which is open to a line-pressure passageway 122.

The manual selector valve 82 comprises an elongate valve chamber 124 which has first, second, third, fourth, fifth and sixth fluid ports 126, 128, 130, 132, 134 and 136, respectively, and a spool valve member 138 which is axially movable in the valve chamber 124. The valve chamber 124 is drained at both ends to the oil reservoir 120. The spool valve member 138 is formed with a first land 140 which is located at one axial end of the valve member and a second land 142 which is located at an intermediate longitudinal portion of the valve member. A first circumferential groove 144 is thus formed between the axially spaced first and second lands 140 and 142 and a second circumferential groove 146 formed in the remaining longitudinal portion of the valve member 138. The first fluid port 126 of the valve chamber 124 is in communication with the pump 80 through the line-pressure passageway 122 and the second, third, fourth, fifth and sixth fluid ports 128, 130, 132, 134 and 136 of the valve chamber 124 lead to fluid passageways 148, 150, 152, 154 and 156, respectively. The spool valve member 138 is mechanically connected to a manual selector lever (not shown) which may be manually operated by a vehicle driver to move the spool valve member 138 to seven different positions which include the parking position P, the reverse position R, the neutral position N, the normal or automatic drive range position D, the manual firstforward-speed-range position 3, the manual secondforward-speed-range position 2 andthe manual firstforward-speed-range position 1, as indicated on a righthand lower part of FIG. 2. The fluid ports of the valve chamber 124 are located in the following manners. When the spool valve member 138 is moved to the parking position P, the first fluid port 126 leading from the line-pressure passageway 122 is closed by the second land 142 of the spool valve member 138 so that no fluid pressure will be admitted into the valve chamber 124 and passed over to the fluid passageways 148, 150,

152, 154 and 156 through the second, third, fourth, fifth and sixth fluid ports 128, 130, 132, 134 and 136, respectively, although the second to sixth fluid ports are open or, more specifically, the second fluid ports 128 is in communication with the first circumferential groove 144 in the valve member 138 and the third, fourth, fifth and sixth fluid ports 130, 132, 134 and 136 are in communication with the second circumferential groove 146 in the valve member 138. The third to sixth fluid ports are drained off through one end of the valve chamber 124. When the spool valve member 138 is axially moved to the reverse position R, the first and second fluid ports 126 and 128 are brought into communication with the first circumferential groove 144 in the spool valve member 138 whereas the third, fourth, fifth and sixth fluid ports 130, 132, 134 and 136 are isolated from the first fluid port 126 by the second land 142 of the spool valve member 138 and are drained off. Fluid communication will thus be established between the line-pressure passageway 122 leading to the first fluid port 126 and the fluid passageway 148 leading from the second fluid port 128. When the spool valve member 138 is axially moved to the neutral position N, then the first fluid port 126 is brought into communication with the first circumferential groove 144 in the valve member 138 but the second fluid port 128 is isolated from the first fluid port 126 by the first land 140 of the valve member 138 and at the same time the third, fourth, fifth and sixth fluid ports 130, 132, 134 and 136 are isolated from the first fluid port 126 by the second land 142 of the valve member 138 and are drained off. No fluid pressure will consequently be passed from the first fluid port 126 over to the fluid passageways 148, 150, 152, 154 and 156 through the second, third, fourth, fifth and sixth fluid ports 128, 130, 132, 134 and 136, respectively, as in the case where the spool valve member 138 is moved to the parking position P. When the spool valve member 138 is axially moved to the normal or automatic drive range position D as illustrated in FIG. 4, both of the first and third fluid ports 126 and 130 are in communication with the first circumferential groove 144 of the valve member 138 whereas the second fluid port 128 is isolated from the first fluid port 126 by means of the first land 140 of the valve member 138 and at the same time and fourth, fifth and sixth fluid ports 132, 134 and 136 are isolated from both of the first and third fluid ports 126 and 130 by the second land 142 of the valve member 138 and are thus drained off. Under these conditions, the fluid pressure in the line-pressure passageway 122 is passed over to the passageway 150 through the first fluid port 126, the first circumferential groove 144 in the valve member 138 and the third fluid port 130. When the spool valve member 138 is axially moved to the manual thirdforward-speedrange position 3, then the first, third and fourth ports 126, 130 and 132 are in communication with the first circumferential groove 144 in the valve member 138 whereas the second fluid port 128 is isolated from the first fluid port 126 by means of the first land 140 of the valve member 138 and at the same time the fifth and sixth fluid ports 134 and 136 are isolated from the first, third and fourth fluid ports 126, 130 and 132 by the second land 142 of the valve member 138 and are drained off. The fluid pressure in the linepressure passageway 122 is consequently passed over to the fluid passageways 150 and 152 through the third and fourth fluid ports 130 and 132, respectively. When sixth fluid port 136 isolated from the first, third, fourth and fifth fluid ports 126, 130, 132 and 134 by the second land 142 of the valve member 138 and is thus drained off. The fluid pressure in the line-pressure passageway 122 is thus directed to the fluid passageway 150, 152 and 154 through the third, fourth and fifth fluid ports 130, 132 and 134, respectively. When the spool valve member 138 is axially moved from the manual second-forward-speed-range position to the manual first-forward-speed-range position 1, then the sixth fluid port 136 is open in addition to the first, third, fourth and fifth fluid ports 126, 130, 132 and 134 which are kept open. The fluid pressure in the line pressure passageway 122 is consequently passed over to the passageways 150, 152, 154 and 156 through the third, fourth, fifth and sixth fluid ports 130, 132, 134 and 136, respectively.

From the above description it will be understood that the line pressure is selectively established in one or more of the five outlet ports 128, 130, 132, 134 and 136 of the manual selector valve 82 in accordance with the following schedules: in the second fluid port 128 when the manual selector valve 82 is in the reverse position R; in the third fluid port 130 when the manual selector valve 82 is in the normal or automatic drive range D or any of the third, secondand firstforwardspeed-range positions 3, 2 and 1; in the third fluid port 132 when the manual selector valve 82 is in any of the manual third-, second-, firstforward-speed-range positions 3, 2 and 1; in the fifth fluid port 134 when the manual selector valve 82 is in either of the manual secondand first-forward-speed-range positions 2 and 1; and in the sixth fluid port 136 when the manual selector valve 82 is in the manual first-forward-speed-range position 1. When, furthermore, the manual selector valve 82 is in any of the positions other than the manual third, secondand first-forward-speed-range positions 3, 2 and 1, the fourth, fifth and sixth fluid ports 132, 134 and 136 are drained off, viz., in communication with the previously mentioned oil reservoir 120 so that no fluid pressure will be developed in the fluid passageways 152, 154 and 156. When the manual selector valve 82 is in the manual third-forward-speed-range po sition 3, the fifth and sixth fluid ports 134 and 136 are drained off so that no fluid pressure will be developed in the fluid passageways 154 and 156. When the manual selector valve 82 is in the manual second-forwardspeed-range position 2, the sixth fluid port 136 is drained off so that no fluid pressure will be developed in the fluid passageway 156. When the manual selector valve 82 is in the manual first-forward-speed-range position, none of the fourth, fifth and sixth fluid ports 132, 134 and 136 is drained off.

The fluid passageway 148 leading from the second fluid port 128 is in communication through a shuttle valve 155 to the apply servo mechanism of the front clutch 18 and to the release side of the servo mechanism of the band brake 35 and through a shuttle valve 157 to the apply servo mechanism of the low-andreverse brake 28. The fluid passageway leading from the third fluid port 130 of the selector valve 82 is in communication with the apply servo mechanism of the rear clutch 20. When, thus, the manual selector valve 82 is moved to the reverse position R with the re sult that the second fluid port 128 is brought into communication with the line-pressure passageway 122, the line pressure is directed to the fluid passageway 148 so that the front clutch 18 is coupled, the lowand-reverse brake 28 is actuated and the band brake 34 is released. When, on the other hand, the manual selector valve 82 is moved to the normal or automatic drive range position D or any one of the manual third-, second and firstforward-speed-range positions 3, 2 and 1 and consequently the third fluid port 130 of the manual selector valve 82 is brought into communication with the linepressure passageway 122, then the line pressure is di rected to the fluid passageway 150 so that the rear clutch 20 is coupled. Operations of the shuttle valves 155 and 157 which are disposed in the fluid passageway 148 will be described later.

Table 2 shows the conditions of the fluid ports 126, 128, 130, 132, 134 and 136 of the manual selector valve 82 as achieved when the selector valve is in the various positions above mentioned, wherein a symbol indicates the open condition of the first fluid port 126 or the condition in which the fluid port other than the first port is in communication with the first port and a symbol x indicates the closed condition of the first fluid port or the condition in which the fluid port other than the first port is isolated from the first port.

The line pressure in the line-pressure passageway 122 is directed to the vacuum-operated throttle valve 84. The vacuum-operated throttle valve 84 cooperates with the pressure modifier valve 86 and is operative to regulate the line pressure to be applied to the clutches and brakes of the power transmission mechanism in relation to the vacuum in the intake manifold of the engine so as to change the gear ratio shift point in accordance with variation in the engine load. The line pressure thus regulated by the combination of the throttle valve 84 and the pressure modifier valve 86 will be herein termed a throttle pressure. The throttle valve 84 comprises an elongate valve chamber 158 which has first, second, third and fourth fluid ports 160, 162, 164 and 166 and a spool valve member 168 which is axially movable in the valve chamber 158. The spool valve member 168 is formed with first, second and third lands 170, 172 and 174 which are axially spaced apart from each other and which thus define a first circumferential groove 176 between the first and second lands 170 and 172 and a second circumferential groove 178 between the second and third lands 172 and 174, as shown. The first and second lands 170 and 172 have equal cross sectional areas S and the third land 174 has a cross sectional areas S which is smaller than the cross sectional areas S of the lands 170 and 172. The first fluid ports 160 is in communication with the linepressure passageway 122 and is so located as to be covered or uncovered by the second land 172 of the spool valve member 168 depending upon the axial relative position of the valve member relative to the port 160. The second fluid port 162 is in communication with a throttle-pressure passageway 180 and is kept open to the first circumferential groove 176 between the first and second lands and 172 irrespective of the axial relative position of the spool valve member 168 in the valve chamber 158. The third fluid port 164 is in communication with a fluid passageway 182 through a restriction or orifice 184 and is kept open to the second circumferential groove 178 between the second and third lands 172 and 174 of the spool valve member 168. The fourth fluid port 166 is a drain port which leads to the oil reservoir l20and which is so located as to be covered or uncovered by the first land 170 of the spool valve member 168 depending upon the axial relative position of the spool valve member 168 in the valve chamber 158.

The pressure modifier valve 86 cooperating with the throttle valve 84 above described comprises an elongate valve chamber 186 which has first, second and third fluid ports 188, 190 and 192. The valve chamber 186 of the pressure modifier valve 86 is conjoined at one end to the valve chamber 158 of the throttle valve 84 through an intermediate port 194 which is in constant communication with the throttle-pressure pas sageway through a restriction or orifice 196. The pressure modifier valve 86 further comprises a spool valve member 198 which is formed with a first land 200 located adjacent to the intermediate port 194 between the valve chambers 158 and 186 of the throttle valve 84 and the pressure modifier valve 86, respectively, and a second land 202 which is axially spaced apart from the first land 200 for defining a circumferential groove 204 between the first and second lands 200 and 202. The first land 200 has a cross sectional area which is equal to the cross sectional area S of the third land 174 of the spool valve member 168 of the throttle valve 84 and the second land 202 has a cross sectional area 5;, which is larger than the cross sectional area S of the first land 200. The first fluid port 188 is in communication with the throttle-pressure passageway 180 and is so located as to be covered or uncovered by the first land 198 of the spool valve member 198 depending upon the axial relative position of the valve member in the valve chamber 186. The second fluid port 190 is in communication with the fluid passageway 182 leading from the third fluid port 178 of the valve chamber 158 of the throttle valve 84 and is kept open to the circumferential groove 204 between the first and second lands 200 and 202 irrespective of the axial relative position of the spool valve member 198 in the valve chamber 186. The third fluid port 192 is a drain port which is in communication with the oil reservoir 120 and which is so located as to be covered or uncovered by the second land 202 of the valve member 198 depending upon the axial relative position of the valve member in the valve chamber 186. The spool valve member 198 is biased to axially move toward the intermediate port 194 by means of a preload spring 206 which is shown as seated at one end on 'the second land 202 of the valve member. The force thus acting on the spool valve member 198 from the preload spring 206 is herein denoted by PS.

As previously noted, the throttle valve 84 is responsive to the vacuum which is developed in the intake manifold of the engine. The throttle valve 84 is thus provided with a vacuum-operated valve actuator which is generally designated by reference numeral 208. The vacuum-operated valve actuator 208 comprises a casing 210 which is divided by a diaphragm member 212 into an air chamber 214 and a vacuum chamber 216 which is in constant communication with the intake manifold of the engine though not shown in FIG. 4. The spool valve member 168 of the throttle valve 84 has an axial extension 218 which projects from the first land 170 of the valve member 168 into the air chamber 214 of the valve actuator 208 and which is connected to the diaphragm member 212 of the valve actuator as shown. The vacuum chamber 216 has accommodated therein a preload spring 220 which urges the diaphragm member 212 toward the valve chamber 158 of the throttle valve 84. The vacuum drawn from the intake manifold of the engine acts on the diaphragm member 212 so that the diaphragm member 212 is moved away from the valve chamber 158 if the force of the vacuum overcomes the opposing force of the preload spring 220 so that the spool valve member 168 of the throttle valve 84 is axially moved away from the intermediate port 194 between the valve chambers 158 and 186 of the throttle valve 84 and the pressure modifier valve 86. The force thus effective to move the spool valve member 168 toward the intermediate port 194 between the valve chambers 158 and 186 is herein represented by Fv. The force Fv is apparently a difference between the force of the vacuum which acts on the diaphragm member 212 and the opposing force which is exerted by the preload spring 220. If the force Fv of the vacuum drawn into the vacuum chamber 216 of the valve actuator 208 yields to the opposing force Fs of the preload spring 220, then the diaphragm member 212 will be moved toward the valve chamber 158 of the throttle valve 84 so that the spool valve member 168 of the throttle valve 84 is moved toward the intermediate port 194 between the aligned valve chambers 158 and 186.

With the throttle valve 84 and the cooperating pres sure modifier valve 86 thus constructed and arranged, the fluid pressure in the throttle-pressure passageway 180 is developed in the first port 160 and accordingly in the first circumferential groove 176 of the spool valve member 168 of the throttle valve 84 and through the orifice 196 in the intermediate port 194 between the valve chambers 158 and 186 of the throttle valve 84 and the pressure modifier valve 86. The fluid pressure thus developed in the first circumferential groove 176 of the spool valve member 168 acts on annular end faces of the first and second lands 170 and 172 of the spool valve member 168. The forces thus exerted on the annular end faces of the first and second lands 170 and 172 of the spool valve member 168 are, however, cancelled by each other because the first and second lands 172 and 174 have the same cross sectional areas S as previously mentioned. The fluid pressure which is developed in the intermediate port 194 between the valve chambers 158 and 186 of the throttle valve 84 and the pressure modifier valve 86 acts on an end face of the third land 174 of the spool valve member 168 of the throttle valve 84 and on an end face of the first land 200 of the spool valve member 198 of the pressure modifier valve 86. The fluid pressure thus acting on the lands 174 and 200 of the spool valve members 168 and 198 urges the spool valve members 168 and 198 axially away from the intermediate port 194 by an equal force because of the fact that the lands 174 and 200 have the same cross sectional areas S as previously mentioned. On the other hand, the fluid pressure which obtains in the fluid passageway 182 is directed through the orifice 5 184 and the port 164 into the second circumferential groove 178 in the spool valve member 168 of the throttle valve 84 and through the port 190 into the circumferential groove 204 in the spool valve member 198 of the pressure modifier valve 86. The fluid pressure directed into the second circumferential groove 178 in the spool valve member 168 of the throttle valve 84 acts on annular end faces of the second and third lands 172 and 174 of the spool valve member 168 and, because of the face that the second land 172 is larger in cross sectional area than the third land 174, urges the spool valve member 168 axially away from the intermediate port 194 between the valve chambers 158 and 186 of the throttle valve 84 and the pressure modifier valve 86. The fluid pressure directed from the fluid passageway 182 into the circumferential groove 204 in the spool valve member 198 of the pressure modifier valve 86 acts on annular end faces of the first and second lands 200 and 202 of the spool valve member 198 and thus urges the spool valve member 198 axially away from the intermediate port 196 because of the larger cross sectional area 8:; of the second land 202 than the cross sectional area S of the first land 200. The face thus urging the spool valve member 198 is opposed by the force of the preload spring 206 which is constantly operative to urge the spool valve member 198 axially toward the intermediate port 194.

When, now, the spool valve member 198 of the pressure modifier valve 86 is axially so positioned by means of the preload spring 206 as to have the first land 200 located to uncover the first fluid port 188 of the pressure modifier valve 86, then the fluid pressure in the throttle-pressure passageway 180 will be directed through the fluid port 162 into the first circumferential groove 176 in the spool valve member 168 of the throttle valve 84, through the orifice 196 into the intermediate port 194 between the valve chambers 158 and 186 of the throttle valve 84 and the pressure modifier valve 86, through the fluid port 188 into the circumferential groove 204 in the spool valve member 198 of the pressure modifier valve 86, and through the fluid port 190, the passageway 182, the orifice 184 and the fluid port 164 into the second circumferential groove 178 of the spool valve member 168 of the throttle valve 84. If the fluid pressure developed in the throttle pressure passageway 180 under these conditions is denoted by P then the spool valve member 168 of the throttle valve 84 will be held in an axial equilibrium position when the following relation is established:

so that (1) lf, thus, the value of F is increased progressively, then the value of P, will increase in direct proportion to the value of F... This means that the throttle pressure P, 0btaining in the passageway 180 when the first fluid port 188 of the pressure modifier valve 86 is uncovered by the first land 200 of the spool valve member 198 is increased as the force F,- exerted on the spool valve member 168 of the throttle valve 84 from the vacuumoperated valve actuator 208 increases. (The increase of the force F,- results from a decrease in the vacuum acting on the diaphragm member 212 of the valve actuator 208 and, in turn, the decrease in the vacuum results from an increase in the engine load.) The thus increased throttle pressure acts on the differential cross sectional areas of the first and second lands 200 and 202 of the spool valve member 198 and urges the spool valve member 198 axially away from the intermediate port 194 between the valve chambers 158 and 186 of the throttle valve 84 and the pressure modifier valve 86. The spool valve member 198 of the pressure modifier valve 86 will be held in a balanced axial position when the following relation is achieved:

when the throttle pressure 1 is increased to a certain level, then the spool valve member 198 of the pressure modifier valve 86 will be axially moved away from the intermediate port 194 and will assume such a position as to have the first land 200 located to cover the fluid port 188 and to have the second land 202 located to be on the point of uncovering the drain port 192 as illustratcd in FIG. 4. Under these conditions, the spool valve member 198 of the pressure modifier valve 86 will become balanced when the combined forces of the throttle pressure acting on the first land 200 of the valve member 198 from the intermediate port 194 and the fluid pressure acting on the differential pressureacting areas of the first and second lands 200 and 202 of the valve member 198 are equalized with the opposing force which is exerted from the preload spring 206. It therefore follows that the pressure in the circumferential groove 204 in the spool valve member 198 decreases as the throttle pressure in the intermediate port 194 increases. If, in this instance, the fluid pressures in the passageways 180 and 182 under such conditions are denoted by P and P,,,, respectively, the spool valve member 168 of the throttle valve 84 will be held in an equilibrium position when the following relation is established:

so that P, S (2) When the force F exerted from the valve actuator 208 is further increased, the throttle pressure P acting on the spool valve members 168 and 198 of the throttle valve 84 and the pressure modifier valve 86 is also increased and reaches a certain level at which the balanced conditions of the spool valve members 168 and 198 will be destroyed. The spool valve member 198 of the pressure modifier valve 86 is consequently further axially moved away from the intermediate port 194 against the opposing force of the preload spring 206 so that the second land 202 of the spool valve 198 of the pressure modifier valve 86 overruns the drain port 192. The fluid passageway 182 is thus brought into communication with the drain port 192 through the fluid port 190 and the circumferential groove 204 in the spool valve 198 of the pressure modifier valve 198 with the result that no fluid pressure obtains in the passageway 182. Under these conditions, the spool valve member 168 of the throttle valve 86 will be held in a balanced position when the force exerted on the spool valve member 168 by the throttle pressure in the intermediate port 194 is equalized with the opposing force F which is imparted to the spool valve member from the vacuumoperated valve actuator 208. If the throttle pressure developed in the throttle-pressure passageway 180 under these conditions is denoted by P then the following relation will be established:

From comparison between Eqs. (1), (2) and (3) it will be understoodthat the throttle pressure increases at stepwise varying rates as the force F from the valve actuator 208 is increased at a constant rate. As will be more clearly seen from the graph of FIG. 3, the rate of increase of the throttle pressure P given by Eq. (2) is higher than the rate of increase of the throttle pressure P given by Eq. (1) and the rate of increase of the throttle pressure P given by Eq. (3) is lower than the rate increase of the throttle pressure P and higher than the rate of increase of the throttle pressure P This will mean that, as the vacuum in the intake manifold of the engine is increased at a constant rate, the throttle pressure increases at a relatively low rate when the manifold vacuum is higher than a certain level of V (so that the force F,. exerted by the valve actuator 208 is smaller than a certain value) and a relatively high rate when the manifold vacuum is at relatively low levels which is higher than a certain level of V When the manifold vacuum decreases from the level V so that the force F from the valve actuator 208 increases beyond a value providing the condition represented by Eq. (3), then the throttle pressure increases at a reduced rate. Such a tendency of the throttle pressure may be adjusted through selection of the dimensional relations of the lands constituting the spool valve members 168 and 198 and the preload springs 208 and 220.

The throttle pressure thus produced by the throttle valve 84 and the pressure modifier valve 86 is passed over to the throttle back-up valve 88 and to the shift valves 108, 110 and 112 through the throttle pressure passageway 180.

The throttle back-up valve 88 is operable to vary the throttle pressure in relation to the line pressure when the manual selector valve 82 is in the manual third-, secondor first-forward-speed-range position or the position 3, 2 or 1 so that the fourth fluid port 132 of the manual selector valve 82 leading to the fluid passage way 152 is in communication with the line-pressure passageway 122. When the manual selector valve 82 is in any of the remaining positions, viz., the parking position P, the reverse position R, the neutral position N or the normal drive position D, the throttle back-up valve 88 passes the throttle pressure without modification. The throttle back-up valve 88 comprises an elongate valve chamber 222 which has first, second, third, fourth and fifth fluid ports 224, 226, 228, 230 and 232 and a spool valve member 234 which is axially movable in the valve chamber 222. The spool valve member 234 is formed with axially spaced first and second lands 236 and 238 having equal cross sectional areas and a circumferential groove 240 which is located between the first and second lands 236 and 238. The first fluid port 224 is in communication with the fluid passageway 152 leading from the fourth fluid port 132 of the manual selector valve 82 and is so located as to have the fluid pressure in the fluid passageway 152 directed onto an axial end face of the first land 236 of the spool valve member 234 so that the line pressure acts thereupon when the fourth fluid port 132 of the manual selector valve 82 is in communication with the line-pressure passageway 122. The second fluid port 226 is also in communication with the fluid passageway 152 and is so located as to be covered or uncovered by the second land 238 of the spool valve member 234 depending upon the axial relative position of the valve member 234 in the valve chamber 222. The third fluid port 228 is in communication with the throttle-pressure passageway 180 and is so located as to be covered by the other axial end portion of the first land 236 of the spool valve member 234 or open to the circumferential groove 240 in the valve member 234 depending upon the axial relative position of the valve member in the valve chamber 222. The fourth and fifth fluid ports 230 and 232 are in communication with a common fluid passageway 242. The fourth fluid port 230 is located to be usually open to the circumferential groove 240 in the spool valve member 234 irrespective of the axial relative position of the spool valve member 234 in the valve chamber 222. The throttle-pressure passageway 180 is thus brought into communication with the fluid passageway 242 through the third fluid port 228, the circumferential groove 240 in the spool valve member 234 and the fourth fluid port 230 when the first land 236 of the valve member 234 is in an axial position to uncover the third fluid port 228. The fifth fluid port 232 communicates with the above-mentioned fluid passageway 242 through a restriction or orifice 244 and is constantly held open. The spool valve member 234 is biased by means of a preload spring 246 toward an axial position to close the second fluid port 226 by the second land 238 thereof and to have the third fluid port 228 open to the circumferential groove 240 therein. The preload spring 246 is shown to be seated on an end face of the second land 238 of the spool valve member 234. The first and second lands 236 and 238 of the spool valve member 234 have equal cross sectional areas so that the forces acting on these lands from the fluid pressure developed in the circumferential groove 240 in the valve member 234 will be cancelled by each other.

When, now, the manual selector valve 82 is in a position other than the manual third-, secondand firstforwardspeed-range positions 3, 2 and 1, then the fourth, fifth and sixth fluid ports 132, 134 and 136 of the manual selector valve 82 are drained off so that no fluid pressure obtains in the fluid passageway 152 leading to the first and second fluid ports 224 and 226 of the throttle back-up valve 88. The spool valve member 234 of the throttle back-up valve 88 is consequently axially moved by the force of the preload spring 246 and a fluid pressure acting on the second land 238 of the spool valve member 234 from the fifth fluid port 232 into the position to close the second fluid port 226 by the second land 238 and to open the third fluid port 228, as illustrated in FlG. 4. The throttle pressure in the throttle-pressure passageway 180 is therefore passed through the throttle back-up valve 88 without modification by the line pressure and is thus directed to the fluid pressure passageway 242 through the third fluid port 228, the circumferential groove 240 in the spool valve member 234 and the fourth fluid port 230. When, however, the manual selector valve 82 is in the third-, secondor first-forward-speed-range position 3, 2 or 1, then the fluid passageway 152 is in communication with the line-pressure passageway 122 through the fourth and first fluid ports 132 and 126 of the manual selector valve 82 so that the line pressure is directed through the passageway 152 to the first and second fluid ports 224 and 226 of the throttle back-up valve 88. The second fluid port 226 being kept closed by means of the second land 238, the spool valve member 234 is moved from the above-mentioned axial position by the line pressure acting on an end face of the first land 236 of the valve member. The third fluid port 228 is consequently covered by the first land 236 of the spool valve member 234 and at the same time the second fluid port 226 is about to be uncovered by the second land 236 of the valve member. The spool valve member 234 will therefore be brought into a balanced position when the combined forces of the preload spring 246 and the fluid pressure acting on the second land 238 of the valve member are equalized with the force resulting from the line pressure which acts on the first land 236 of the valve member from the first fluid port 224. The fluid pressure in the passageway 242, which is now disconnected from the throttle-pressure passageway 180 with the third fluid port 228 kept closed by the land 236 of the spool valve member 234, is related to the line pressure and is lower than the line pressure by a value which is equal to the force of the preload spring 246. Thus, the fluid pressure in the linepressure passageway 122 will vary without respect to the throttle pressure during the manual third-, secondor first-forward-speed-range condition and will be varied in relation to the throttle pressure during the remaining operational conditions of the transmission. The force of the preload spring 246 is usually so selected as to be an extremely small valve so that the fluid pressure developed in the passageway 242 during the manual third-, secondand first-forward-speed-range conditions is approximately equal to the line pressure. The fluid pressure thus delivered to the passageway 242 is distributed to the control pressure regulator valve and the idle valve 92.

The control pressure regulator valve 90 is adapted to regulate the line pressure and comprises an elongate valve chamber 248 which has first, second, third, fourth, fifth, sixth and seventh fluid ports 250, 252, 254, 256, 258, 260 and 262, respectively, and first and second spool valve members 264 and 266 which are axially movable in the valve chamber 248. The first spool valve member 264 is formed with axially spaced first, second, third and fourth lands 268, 270, 272 and 274, a first circumferential groove 276 located between the first and second lands 268 and 270, a second circumferential groove 278 located between the second and third lands 270 and 272, a third circumferential groove 280 located between the third and fourth lands 272 and 274, and an axial extension 282 projecting from the fourth land 274. The second, third and fourth lands 270, 272 and 274 have substantially equal cross sectional areas which are larger than the cross sectional area of the first land 268. The second spool valve member 266 is formed with axially spaced first and second lands 284 and 286, a circumferential groove 288 located between the lands 284 and 286, and an axial extension 290 which projects from the second land 286 toward the axial extension 282 of the first spool valve member 264. The first land 284 of the spool valve member 266 is smaller in cross sectional area than the second land 286. Of the first to seventh fluid ports above-mentioned, the first to fifth fluid ports are located in association with the first spool valve member 264 and the sixth and seventh fluid ports are located in association with the second spool valve member 266. The first fluid port 250 is in communication with the line-pressure passageway 122 through a restriction or orifice 292 and is located to have the line pressure in the passageway 122 directed onto the axial end face of the first land 268 of the first spool valve member 264. The second fluid port 252 is located to provide constant fluid communication between the line-pressure passageway 122 and the third circumferential groove 280 in the first spool valve member 264 so that the line pressure constantly acts on the confronting annular end faces of the third and fourth lands 272 and 274 of the first spool valve member 264. The forces thus exerted on the third and fourth lands 272 and 274 of the first spool valve member 264 by the line pressure directed to the third circumferential groove 280 from the second fluid port 252 are cancelled by each other because the third and fourth lands 272 and 274 have substantially equal pressure acting areas. The third fluid port 254 is in communication through a shuttle valve 294 to the fluid passageway 150 leading from the third fluid port 130 of the manual selector valve 82 and is so located as to be open to the first circumferential groove 276 in the first spool valve member 264 or closed by the second land 70 of the valve member 264 depending upon the axial relative position of the first spool valve member 264 in the valve chamber 248. The snuttle valve 290 is adapted to pass the fluid pressure from the passageway 150 to the third fluid port 254 at all times when the line pressure is developed in the passageway 150 as will be discussed again as the description proceeds. The fourth fluid port 256 is a drain port which is in communication with the oil reservoir 120 and which is constantly open to the second circumferential groove 278 in the first spool valve member 264 irrespective of the axial relative position of the first spool valve member 264 in the valve chamber 248. The fourth fluid port 256 may be brought into communication with the second fluid port 252 when the spool valve member 264 is axially moved so that the third land 272 thereof overruns the second fluid port 252 from the shown position. The fifth fluid port 258 is in communication with the torque converter (FIG. 1) and is so located as to be covered or uncovered by the fourth land 274 of the first spool valve member 264. When uncovered by the fourth land 274 of the spool valve member 264, the fifth fluid port 258 is in communication with the second fluid port 252 through the third circumferential groove 280 in the valve member 264 so that the line pressure is supplied to the torque converter through the fifth fluid port 258. The first spool valve member 264 is biased by a preload spring 296 to move toward an axial position closing the fifth fluid port 258, opening the third fluid port 254, and blocking the fluid communication between the second and fourth fluid ports 252 and 256, as shown by the third land 272 of the valve member 264. Of the sixth and seventh fluid ports 260 and 262 which are associated with the second spool valve member 266, the sixth fluid port 260 is in communication with the fluid passageway 242 leading from the fifth fluid port 232 of the throttle back-up valve 88 and is so located as to have the fluid pressure directed from the fluid passageway 242 onto the axial end face of the first land 284 of the second spool valve member 266, which is consequently is urged toward the first spool valve member 264. The

seventh port 262 is in constant communication with the fluid passageway 150 leading from the third fluid port of the manual selector valve 82 and is so located as to be covered or uncovered by the second land 286 of the second spool valve member 266 as the spool valve member 266 is axially moved in the valve chamber 248.

When the manual selector valve 82 is in the normal drive range position D and or in anyone of the manual third-, secondand first-forward-speed-range positions 3, 2 and 1, the line pressure is developed in the third fluid port 130 of the manual selector valve 82. The line pressure is thus directed through the passageway to the seventh fluid port 262 of the control pressure regulator valve 90. The line pressure acts on the differential pressure acting areas of the first and second lands 284 and 286 of the second spool valve member 266 and, combined with the fluid pressure acting on the end face of the first land 284 from the sixth fluid port 260 leading from the throttle back-up valve 88 through the passageway 242, causes the second spool valve member 266 to move toward the first spool valve member 264 until the axial extension 290 of the second spool valve member 266 is brought into abutting engagement with the axial extension 282 of the first spool valve member 264. When, on the other hand, the manual selector valve 82 is in any one of the positions other than the positions D, 3, 2 and l, the third fluid port 130 of the selector valve 82 is drained off so that no fluid pressure is developed in the fluid passageway 150 (in which conditions the shuttle valve 290 is in a condition blocking the fluid communication between the passageway 150 and the third fluid port 254 as will be discussed later). Under these conditions, the fluid pressure directed to the sixth fluid port 260 from the passageway 242 acts on the end face of the first land 284 of the second spool valve member 266, which is consequently axially moved toward the first spool valve member 264 until the axial extension 290 of the second spool valve member 266 is brought into abutting engagement with the axial extension 282 of the first spool valve member 264. In whichsoever position the manual selector valve 82 may be held, the second spool valve member 266 of the pressure regulator valve 90 is thus in engagement with the first spool valve member 264 and, as a consequence, the first and second spool valve members 264 and 266 are axially movable as a unit in the valve chamber 248. The first and second spool valve members 264 and 266 will therefore be held in balanced positions when the combined forces resulting from the line pressure acting on the end face of the first land 268 of the first spool valve member 264 from the first fluid port 250 and the fluid pressure which may act on the differential working faces of the first and second lands 268 and 270 of the first spool valve member 264 from the third fluid port 254 are equalized with the combined forces resulting from the fluid pressure acting on the end face of the first land 284 of the second spool valve member 266 from the sixth fluid port 260, the line pressure which may act on the differential working faces of the first and second lands 284 and 286 of the second spool valve member 266 from the seventh fluid port 262 and the force of the preload spring 296. When the manual selector valve 82 is in any of the manual third-, secondand first-forward-speed-range positions 3, 2 and l, the fluid pressure directed from the throttle back-up valve 88 to the sixth fluid port 260 through the passageway 242 is substantially equal to the line pressure as previously mentioned. During any of these conditions, therefore, the regulator valve 90 is operative to regulate the line pressure without respect to the vacuum in the intake manifold of the engine. When, however, the manual selector valve 82 is in any of the positions other than the manual third-, secondand firstforward-speed-range positions 3, 2 and 1, the throttle pressure delivered from the throttle valve 84 and the pressure modifier valve 86 is directed through the throttle back-up valve 88 to the sixth fluid port 260 of the regulator valve 90, which is therefore operative to regulate the line pressure in relation to the throttle pressure or, in other words, to the vacuum developed in the intake manifold of the engine. During the normal drive range condition or any of the manual third-, second-and first-forward-speed-range conditions of the transmission, furthermore, the fluid pressure to act on the differential working areas of the first and second lands 268 and 270 of the first spool valve member 264 is the line pressure which is directed from the passageway 150 to the third fluid port 254 through the shuttle valve 294. When the manual selector valve 82 is in any of the positions other than the above-mentioned positions, viz., in the parking position P, the reverse position R or the neutral position N, the fluid passageway 150 is drained off so that no fluid pressure is developed in the passageway 150 as previously mentioned. If, under these condition, the engine is in an idling condition, another fluid pressure, herein called a hysteresis pressure as will be discussed later, is directed through the shuttle valve 294 to the third fluid port of the pressure regulator valve 90. For this purpose, the shuttle valve 294 is in communication with the previously mentioned idle valve 96 through a fluid passageway 298. The idle valve 96 operates on the hysteresis pressure which is supplied thereto from the hysteresis valve 92 through the solenoid-operated control valve 94.

The hysteresis valve 92 comprises an elongate valve member 300 which has first, second, third, fourth and fifth fluid ports 302, 304, 306, 308 and 310 and a spool valve member 312 which is axially movable in the valve chamber 300. The spool valve member 312 is formed with axially spaced first and second lands 314 and 316 and a circumferential groove 318 which is located between the lands 314 and 316. The first fluid port 302 is in communication with the line-pressure passageway 122 and is so located as to be covered by the first land 314 of the spool valve member 312 or open to the circumferential groove 318 in the valve member 312 depending upon the relative axial position of the valve member. The second fluid port 304 is in communication with a fluid passageway 320 leading to the sole,- noid-operated control valve 94 and is so located as to be constantly open to the circumferential groove 318 in the spool valve member 312. When the first fluid port 302 is uncovered by the first land 314 of the spool valve member 312, then the line pressure in the first fluid port 302 is thus directed to the passageway 320 through the second fluid port 304. The second fluid port 304 is further in communication through a restriction or orifice 322 with the third fluid port 306, into which does project an axial end portion of the first land 314 of the spool valve member 312. The fourth fluid port 308 is a drain port which is so located as to be covered or uncovered by the second land 316 of the spool tion with the drain port 308. The fifth fluid port 310 is constantly drained off. The spool valve member 312 is biased by a preload spring 324 toward an axial position opening the first and second fluid ports 302 and 304 and concurrently closing the fourth fluid port 308 by the second land 316, as illustrated. The spool valve member 312 will thus be held in a balanced position when the fluid pressure acting on the axial end face of the first land 314 of the valve member 312 from the third fluid port 306 is equalized with the opposing force of the preload spring 324. Under this balanced condition of the spool valve member 312, the first fluid port 302 is on the point of being closed by the first land 314 and the fourth or drain port 308 is on the point of being open to the circumferential groove 318 in the spool valve member 312. In the second fluid port 304 is thus developed a substantially constant fluid pressure which is determined by the force of the preload spring 324. The fluid pressure thus developed is the hysteresis pressure previously mentioned. The hysteresis pressure is fed to the solenoid-operated control valve 94 en route the fluid passageway 320.

The solenoid-operated control valve 94 comprises an elongate valve chamber 326 which has first, second, third, fourth and fifth fluid ports 328, 330, 332, 334 and 336, respectively, and a spool valve member 338 which is axially movable in the valve chamber 326. The spool valve member 338 is formed with axially spaced first and second lands 340 and 342 and a circumferential groove 344 which is located between the lands 340 and 342. The first fluid port 328 is in communication through a restriction or orifice 346 with the fluid passageway 320 leading from the second fluid port 304 of the hysteresis valve 92 and is so located as to have the fluid pressure directed from the passageway 320 onto the axial end face of the first land 340 of the spool valve member 338, which is therefore urged to axially move away from the fluid port 328. The second fluid port 330 is in communication with the fluid passageway 320 and according through the restriction 346 with the first fluid port 328 and is so located as to be covered by the first land 340 of the spool valve member 338 or open to the circumferential groove 344 in the valve member 338 depending upon the axial relative position of the valve member in the valve chamber 326. The third fluid port 332 is in communication on one hand with a fluid passageway 348 leading to the idle valve 96 and on the other hand with a fluid passageway 350 leading to the kick-down valve 98. The third fluid port 332 is located to be constantly open to the circumferential groove member 338 is biased by a preload spring 252 toward an axial position to have the first land 340 located to uncover the second fluid port 330 and to axial end face of the first land 340 of the spool valve member 338 is formed an orifice 354 which is in communication at one end with the first fluid port 328 and at the other end with the previously mentioned oil reservoir 120. A solenoid-operated valve actuator 356 has a plunger 358 which projects into the orifice 354. The solenoidoperated valve actuator 356 is electrically connected to a power source (not shown) over a switch 360 which is responsive to the movement of an accelerator pedal (not shown) of the vehicle. The switch 360 is herein assumed to be closed when the accelerator pedal is either released or depressed all the way down, viz., in response to idling or kick-down conditions of the engine. When the switch 360 is open and accordingly the solenoid-operated valve actuator 356 remains deenergized, the plunger 358 is held in a position projecting into the orifice 354. The orifice 354 is consequently closed by the plunger 358 so that a fluid pressure (which is the hysteresis pressure directed from the second fluid port 304 of the hysteresis valve 92 through the fluid passageway 320) is developed in the first fluid port 328. The spool valve member 338 is thus axially moved away from the first fluid port 328 toward a position to close the second fluid port 330 by the second land 340 of the valve member 338 and to provide communication between the third and fourth fluid ports 332 and 334 through the circumferential groove 344 in the valve member 338. Under these conditions, the fluid passageway 320 leading from the second fluid port 304 is isolated from the passageways 348 and 350 and the passageways 348 and 350 are drained off through the fourth fluid port 334. When, however, the switch 360 is closed responsive to the idling or kick-down conditions of the engine. The solenoid-operated valve actuator 56 is energized so that the plunger 358 of the valve actuator 356 is withdrawn from the position closing the orifice 354. The first fluid port 328 is consequently drained off through the orifice 354 until the spool valve member 338 is axially moved by the force of the preload spring 352 to a position closing the orifice at the axial end face of the first land 340 of the spool valve member 338 as shown. Under these conditions, fluid communication is established between the second and third fluid ports 330 and 332 through the circumferential groove 344 in the spool valve member 338 and at the same time the fourth fluid port 334 is closed by the second land 342 of the spool valve member 338. The hysteresis pressure in the passageway 320 leading from the second fluid port 304 of the hysteresis valve 92 is therefore passed over to the fluid passageways 348 and 350 through the second and third fluid ports 330 and 332 of the solenoid operated control valve 94. The hysteresis pressure developed by the hysteresis valve 92 is in this manner delivered to the idle valve 96 through the fluid passageway 348 and to the kick-down valve 98 through the passageway 350.

The idle valve 96 comprises an elongate valve chamber 362 which has first, second, third, fourth and fifth fluid ports 364, 366, 368, 370 and 372, respectively, and a spool valve member 374 which is axially movable in the valve chamber 362. The spool valve member 374 is formed with axially spaced first and second lands 376 and 378 having substantially equal cross sectional areas and a circumferential groove 380 which is located between the first and second lands 376 and 378. The first fluid port 364 is in communication with the fluid passageway 242 leading from the fifth fluid port 232 of the previously described throttle back-up valve 88 and is so located as to have the fluid pressure directed from the passageway 242 onto the axial end face of the first land 376 of the spool valve member 374 for urging the spool valve member 374 axially away from the first fluid port 364. The second fluid port 366 is in communication with the fluid passageway 348 leading from the third fluid port 332 of the solenoid-operated control valve 94 and is so located as to be covered by the first land 376 of the spool valve member 374 or open to the circumferential groove 380 in the spool valve member 374 depending the axial relative position of the valve member 374 in the valve chamber 362. The third fluid port 368 is in communication on one side with the previously mentioned fluid passageway 298 leading through the shuttle valve 294 to the third fluid port 254 of the pressure regulator valve and on the other side with a fluid passageway 382 which leads to the downshift valve 106 to be described later. The third fluid port 368 is so located as to be constantly open to the circumferential groove 380 in the spool valve member 374 irrespective of the axial relative position of the spool valve member 374 in the valve chamber 362. The fourth fluid port 370 is a drain port which leads to the oil reservoir and which is so located as to be covered by the second land 378 of the spool valve member 374 when the spool valve member 374 is axially moved in the valve chamber 362 from the shown position. The fifth fluid port 372 is also a drain port which is constantly in communication with the oil reservoir 120. The spool valve member 374 is biased by means of a preload spring 384 toward an axial position in which the first land 376 of the valve member is located to uncover the second fluid port 366 and concurrently the second land 378 of the valve member is located to cover the fourth fluid port 370, as illustrated.

When, as previously discussed, the accelerator pedal is released and as a consequence the valve actuator 356 of the solenoid-operated control valve 94 is energized, the spool valve member 338 of the control valve 94 is axially moved to a position providing communication between the fluid passgeways 320 and 348 through the second and third fluid ports 330 and 332 of the valve 94. The hysteresis pressure delivered from the second fluid port 304 of the hysteresis valve 92 is thus directed to the second fluid port 366 and acts on the opposite end faces of the first and second lands 376 and 378 of the spool valve member 374. The forces thus exerted on the first and second lands 376 and 378 are, however, cancelled by each other because the lands have substantially equal cross sectional areas so that the spool valve member 374 will be axially balanced when the force resulting from the throttle pressure acting on the axial end face of the first land 376 of the spool valve member 374 from the first fluid port 364 is equalized with the opposing force of the preload spring 384. Under the idling condition in which the accelerator pedal is released, however, the throttle pressure delivered from the throttle valve 86 is maintained at a rela tively low level so that the force resulting from the fluid pressure acting on the first land 376 of the spool valve member 374 from the first fluid port 364 will be overcome by the force of the preload spring 384. The spool valve member 374 is therefore axially moved to the 

1. A hydraulic control system of an automatic power transmission for selectively producing a multiplicity of forward-speed gear ratios and a reverse-speed gear ratio, comprising a source of line pressure, a throttle valve responsive to load on an engine for producing from the line pressure a throttle pressure related to the engine load, a governor valve responsive to vehicle speed for producing from the line pressure a governor pressure related to the vehicle speed, a hysteresis valve for producing from the line pressure a substantially constant hysteresis pressure which is higher than said governor pressure, a kick-down valve communicating with said throttle valve and operative to pass therethrough the hysteresis pressure when the throttle pressure is increased to a level substantially representative of kick-down condition of the engine, and a plurality of shift valves each having a downshift position and an upshift position, at least one of said shift valves having first differential pressure-acting areas and second differential pressure-acting areas and formed with a first fluid port to communicate with said source of line pressure during automatic drive range condition of the transmission, the line pressure in the first fluid port acting upon the first differential pressure-acting areas for urging the shift valve toward the downshift position when the shift valve is in the downshift position, the line pressure in the first fluid port being isolated from the first differential pressure-acting areas when the shift valve is in the upshift position, a second fluid port which is in constant communication with said throttle valve, the throttle pressure in the second fluid port urging the shift valve toward the downshift position, a third fluid port communicating with said governor valve, the governor pressure in the third fluid port urging the shift valve toward the upshift position, a fourth fluid port in constant communication with said hysteresis valve and located to be closed when the shift valve is in the upshift position, the hysteresis pressure in the fourth fluid port acting upon said second differential pressure-acting areas for urging the shift valve toward the downshift position when the shift valve is in the downshift position, and a fifth fluid port communicating with said kick-down valve and located to be closed when the shift valve is in the downshift position, the hysteresis pressure passed through the kick-down valve to the fifth fluid port acting upon the second differential pressureacting areas for urging the shift valve toward the downshift position when the shift valve is in the upshift position.
 2. A hydraulic control system as set forth in claim 1, further comprising a control valve responsive to kick-down condition of the engine and connected between said hysteresis valve and said kick-down valve, said control valve being open to pass therethrough the hysteresis pressure from the hysteresis valve to the kick-down valve in response to the kick-down condition of the engine.
 3. A hydraulic control system as set forth in claim 1, further comprising a control valve responsive to kick-down and idling conditions of the engine and connected between said hysteresis valve and said kick-down valve for being open to pass therethrough the hysteresis pressure from the hysteresis valve to the kick-down valve in rEsponse to the kick-down condition of the engine.
 4. A hydraulic control system as set forth in claim 3, further comprising an idle valve connected between the throttle valve and said hysteresis valve for being open to pass therethrough the hysteresis pressure from said control valve in response to the throttle pressure of a level representative of idling condition of the engine, and a two-position valve having a first inlet port communicating with the governor valve, a second inlet port communicating with the idle valve and an outlet port communicating with the third fluid port of the shift valve, said twoposition valve being operative to have the first inlet port open to pass the governor pressure from the governor valve to the third fluid port of the shift valve when the idle valve is in closed condition or to have the second inlet port open to pass the hysteresis pressure from the idle valve to the third fluid port of the shift valve when said control valve and said idle valve are concurrently open.
 5. A hydraulic control system as set forth in claim 4, further comprising a downshift valve connected between said governor valve and said first inlet port of said two-position valve and communicating with said kick-down valve, said downshift valve being operative to boost the governor pressure from the governor valve when the kick-down valve is closed responsive to non-kickdown condition of the engine so that the hysteresis pressure is not directed to the downshift valve. 